JP2002516966A - Vehicle transmission powertrain - Google Patents

Abstract

A power dividing device for a vehicle transmission to provide two rotating outputs having variable relative speed of rotation from a single rotatable input. The device includes a first rotational element and a second rotational element housed within the first element and rotatable about the same axis as the second rotational output. A first fluid chamber associated with the first rotational element and first regulating means to varies the volume of the first chamber in response to rotation of the first rotational element. A second fluid chamber associated with the second rotational element and second regulating means likewise varies the volume to the flow between the first and second chambers the relative timing of variation of the volumes of the first and second chambers determines the speed of the second rotational element in response to rotation of the first rotational element.

Description

DETAILED DESCRIPTION OF THE INVENTION

TECHNICAL FIELD The present invention relates to a vehicle power train.

[0002] Vehicle powertrains typically control a geared transmission with a power source controlled by a manual or series of clutches and brakes connected via a clutch mechanism, both of which typically have three or more different drive ratios. I will provide a. Most vehicles still use internal combustion engines as a power source. Internal combustion engines typically have a narrow range of operating speed (rpm) that matches their torque or drive as "sweet spots" that provide maximum fuel efficiency and minimal pollution.

[0003] The effect of a stepped transmission is that in each gear or gear stage, the engine first experiences a high torque that requires low revolutions, causing the piston to move slower than the flame front of the combustion during the combustion stroke. Is to cause
As the engine speed increases, it passes through the speed of the sweet spot of balanced operation and produces torque from the frame front and the piston moving at the speed of the inflated air. The engine speed then exceeds the balance between speed and torque demand, enters an overspeed state until it engages the next gear, and the above is repeated again, shifting to the next successive gear.

[0004] On a spacious highway, having an overdrive gear provides additional efficiency, and again once again balances engine speed and torque demand. This is usually a fixed gear ratio, which is generally the maximum capacity of most engines for driving flat or moderately undulating terrain. When encountering slopes, most automatic transmissions require the driver driver to operate the manual button or automatically kick down to a lower gear ratio that is relatively large in harmony with the resulting inefficiency. When encountering loads in the mid-top gear / overdrive range, transmissions often cycle up and downshifts in relatively steep cycles, with the result that both drivers and passengers experience apparent acceleration and deceleration.

In order to operate the engine constantly and harmoniously at the sweet spot of the engine speed balanced with the required torque and the piston speed, which are balanced between the flame front of the combustion cycle and the expansion gas, the prior art requires a variable speed input. There are attempts to combine the various gear drive systems required. Generally, friction driven conical pulleys or rollers are used to achieve speed fluctuations by sliding tapered cones relative to one another. The amount of torque that can be transmitted by this means is minimal and the wear resulting from this option results in inoperability. V-shaped pulleys and matching extended V-pulleys using wide V-belts have had limited success in the low horsepower region.

[0006] Variable speed transmissions using prior art hydraulic pumps and motors are designed to increase the flow rate of hydraulic oil as the output speed increases, causing additional frictional losses due to the increased flow rate. . Attempts to minimize this loss have used higher pressures and lower flow rates. This results in very poor low speed torque characteristics. Due to the energy loss, the resulting loss of heat and the efficiency and performance resulting in the need to dissipate cooling through additional circuits and coolers, this type of transmission is not suitable for high speed vehicle applications. It is not feasible, meaning it only applies to low-speed vehicles and mobile machines.

In applicant's International Patent Application No. PCT / AU97 / 00714, a vehicle power transmission that can be continuously controlled over a set operating range by an outer transmission controlled by two rotary inputs and an inner differential assembly. And a powertrain. All inputs and outputs from the microprocessor are continuously monitored, and continuous real-time micro-adjustments ensure easy and smooth drivers, passengers, fuel efficiency and reduced pollution. A single power source can also be used to drive two inputs, for example, via a hydraulic drive system, and in this international application the split engine or two power sources, i.e. one power unit, is mainly used Operating continuously at a sweet spot for maximum fuel and pollution efficiency, a second power source is used to balance the variable input power. In the disclosure of this International Application PCT / AU97 / 00714, which is hereby incorporated by cross-reference and hereinafter referred to as "applicant's earlier application."

DISCLOSURE OF THE INVENTION In order to minimize major tooling costs for high volume automotive manufacturing,
This is advantageous for use in existing high volume engine manufacturing. This is also to achieve the benefits of continuous operation in a sweet spot of maximum fuel efficiency and minimal pollution, with a perfectly balanced single source of power at maximum balanced speed and torque, and to extend the operation to the full drive range of the vehicle. It is desirable to maintain accuracy in

[0009] It is an object of one aspect of the present invention to provide a power split device having two rotational outputs having a variable relative rotational speed from a single rotational input. It is an object of another aspect of the present invention to provide a vehicle powertrain and power transmission that can be continuously controlled over a set operating range.

In a first aspect, the present invention provides a power split device having two rotational outputs having a variable relative rotational speed from a single rotational input, the device being rotationally driven about a rotational axis. A first rotating element having a first rotating output, a second rotating element rotatable about the rotation axis and having a second rotating output, a first fluid chamber associated with the first rotating element; First regulating means for changing the volume of the first chamber in response to rotation of the rotating element, a second fluid chamber associated with the second rotating element, and the second fluid chamber in response to rotation of the second rotating element. Second regulating means for changing the volume of the two chambers, at least regulating the transfer of a closed fluid flow between the first and second chambers during rotation of the first and second rotating elements. Commutator means for properly establishing, wherein the relative timing of the change in volume of the first and second chambers determines the rotational speed of the second rotary element in response to the rotation of the first rotary element.

In another aspect, the present invention provides a power transmission unit that includes a power split device, an outer main transmission, and an inner differential gear assembly, the power split device having two rotational outputs having variable relative rotational speeds, The device comprises:
A first rotating element driven about a rotation axis by the power unit and having a first rotation output, a second rotation element rotatable about the rotation axis and having a second rotation output, and associated with the first rotation element. A first fluid chamber; first regulating means for changing a volume of the first chamber in response to rotation of the first rotating element; a second fluid chamber associated with the second rotating element; Second regulating means for changing the volume of the second chamber in response to rotation of the rotary element, and closed fluid flow communication between the first and second chambers during rotation of the first and second rotary elements Commutator means for establishing at least regularly the relative timing of the change in the volume of the first and second chambers. The main transmission has two rotation input means respectively driven by the first rotation output and the second rotation output of the power split device, and The input means is rotatably connected to the rotary output means, whereby the rotation speed of the output means varies in proportion to the algebraic means of the rotation speed of the two input means, and the differential gear assembly is located inside the main transmission. And rotational input means operably connected to the two differential rotary output means, wherein the output means of the main transmission and the input means of the differential gear assembly are operatively connected.

In another aspect, the present invention provides a continuous control over a set operating range including a single power unit, a power split device, an outer main transmission, and a power transmission unit including an inner differential gear assembly. In a possible vehicle powertrain, a power split device provides two rotational outputs having variable relative rotational speeds, a first rotational element driven about the rotational axis by the power unit and having a first rotational output, and the rotational shaft. A second rotating element rotatable around the first rotating element and having a second rotating output, a first fluid chamber associated with the first rotating element, and varying a volume of the first chamber in response to rotation of the first rotating element. First regulating means and a second flow associated with the second rotating element. A second regulating means for changing the volume of the second chamber in response to the rotation of the second rotating element; and a second regulating means for rotating the first and second rotating elements during rotation of the first and second rotating elements. Commutator means for at least regularly establishing a closed fluid flow transmission, the relative timing of the change in volume of the first and second chambers being dependent on the rotation of the first rotating element. The main transmission has two rotation input means which are respectively driven by the first rotation output and the second rotation output of the power split device, and the two input means serve as the rotation output means. Operably connected so that the rotational speed of the output means varies in proportion to the algebraic means of the rotational speeds of the two input means, The assembly is located inside the main transmission,
A vehicle power train having rotation input means operably connected to two different rotation output means, wherein the output means of the main transmission and the input means of the differential gear assembly are operably connected. provide.

In accordance with the present invention, a single power unit supplies one drive line of a transmission at a constant power source speed and a second drive line from the same power unit with a fluid containment volume or rod, preferably hydraulic fluid. Then, the reaction force without loss of energy while balancing the power of the transmission can act directly on the power source.

[0014] Preferably, the second drive line mechanically changes the amount of trapped oil.
A change in the speed of the two inputs to the mission results in a change in the output speed according to the following equation:
cause. V_out= 2 × V_secondary-V_primary  Where V_outIs the output speed, V_secondaryIs the second rotation of the power split device
The speed of the secondary input provided by the output, V_primaryIs power
The primary input speed provided by the first rotational output of the splitting device.

[0015] The power split device according to the present invention uses a unique method of proportional control of the trapped oil, so that the second rotational output preferably always turns in the same direction as the first rotational output, including forward and reverse. Except when used in a machine requiring the same speed in reverse, in which case the second rotational output can be reduced to zero rpm, three minutes of the speed of the first rotational output when the transmission output is reverse. Does not rotate at less than 1 in the case of neutral dynamic lock, rotates at half the speed of the first rotation output,
At the time of full forward movement, the motor rotates at the same speed as the first rotation output.

[0016] The differential gear assembly is disposed within the main transmission and has rotational input means operably connected to two different rotational output means, wherein the differential means is coupled to the output means of the main transmission and differentially. The input means of the gear assembly is operably connected.

The control means provides and includes means for receiving command input and for determining performance parameters associated with operation of the powertrain. The performance parameters include the power unit load, which is the pressure of the trapped rod of hydraulic oil in the secondary drive line that traps the power of the outer transmission,
The rotational speed of the first primary drive line and the secondary drive line of the main transmission, the load of each of the two output means of the differential gear assembly, and the rotational speed of each of the two output means of the differential gear assembly.

The control means provides a closed loop feedback control that continuously monitors, analyzes and adjusts the performance parameters in response to the command input.

The power unit may be any range of conventional internal combustion engine types-as well as rotary or gas turbine engines, including gasoline (constant cycle) engines, diesel engines, regulated balanced rotary engines with compression and expansion cycles. It can be constructed from conventional electric motor types. Gasoline and diesel internal combustion engines are preferred because mass production technology has been established at relatively low production costs.

Preferably, the rotational power output means of the primary and secondary drive lines rotate in the same direction with respect to the power transmission unit.

Preferably, at least one rotational power output means of the two power drive lines is selectively operably connected to the input means of the main transmission of the power transmission unit by a clutch output. Alternatively, the microprocessor that detects the zero stop input is positioned such that the control mechanism for the rod confined in hydraulic fluid determines the secondary driveline rotational speed at half the rotational speed of the primary drive input in the transmission. Thus, the 2: 1 input ratio outputs a command to cause a zero output rotation by dynamic lock at the zero output position.

The operable connection can also include gears, chains, belts, electric, hydraulic or direct engine drive shaft connection means.

The power unit, primary and secondary drives, outer transmission and internal differential may be conveniently enclosed by a common housing or may be conveniently removable from the power unit as an assembly.

The powertrain of the present invention can have a power unit configuration that is optimized for different applications. For example, a constant-speed electric motor provided so that it can be used as a generator for regenerating braking and deceleration energy. Any type of power unit, operating at a constant maximum efficiency, can be used for renewable and renewable power demands, such as high speed flywheels in direct or vacuum cases or commercial operation providing extremely low pollution hybrids. Either via electromagnetic bearings used to provide the proper braking.

The main transmission of the power transmission unit may be advantageous
And a gear train. Preferably two inputs of the main transmission
The means comprises a first bevel gear and a planetary gear set arranged coaxially to rotate about a first axis.
Consists of assemblies. Preferably, the planetary gear assembly has an axis perpendicular to the first axis.
Annular pinion carrier rotatably carrying an internally disposed planetary umbrella pinion having
Consists of ya. For convenience, the first bevel gear and the annular pinion carrier are single acting.
Primary and secondary drive lines or primary and secondary from power source
Provides power to the transmission unit through both of the driveline
Can operate independently, such as a combined hybrid power source with a single output shaft
It can be connected to a rotary power source. Preferably the output of the main transmission
Means are provided for rotating both the first bevel gear and the planetary gear assembly for rotation about the first axis.
It comprises a second bevel gear arranged coaxially with the yellowtail. Advantageously, the planetary gear asen
The yellowtail has first and second bevel gears both meshing with a planetary bevel pinion gear.
It is arranged between two bevel gears. As mentioned, the main transmission is
, Consisting of a continuously variable transmission, the speed of the output means being
It varies according to the speed of the input means according to the following equation. V_out= 2 × V_secondary-V_primary  Where V_outIs the output speed, V_secondaryIs the speed of the secondary input (
Planetary gear assembly) and V_primaryIs the primary input speed.

Preferably, the input means of the differential gear assembly of the power transmission unit comprises a differential bevel pinion gear disposed radially inward of the main transmission rotating about an axis perpendicular to the first axis. Become. Preferably, the output means of the differential gear assembly comprises two differential bevel side gears coaxially arranged to mesh with a differential pinion gear rotating about a first axis.

[0027] Preferably, the first and second bevel gears of the main transmission are each formed at the center and have holes extending in the axial direction. Preferably, the two differential bevel side gears of the differential gear assembly extend axially outward through holes in the first and second bevel gears of the main transmission and are preferably operably connectable to a drive wheel. Like a half axle, two coaxially aligned power output members are mounted centrally on opposite ends.

Preferably, the main transmission is operably connected to the differential gear assembly by a differential frame connected to the second bevel gear of the main transmission, the differential frame being connected to the differential gear assembly. Carry a differential pinion gear. As already mentioned, the differential gear assembly has the functionality of a conventional automotive differential gear. This configuration provides a tuned rotary engine or turbine in the form of a "pancake" with a hollow shaft to control reactive energy between a common power unit and primary and secondary drive lines from a single or hybrid power supply or It is particularly advantageous to use primary and secondary power units and / or drive lines such as primary and secondary drive lines with hollow shafts consisting of gears, sprockets or axial and radial pistons, hydraulic restraints and power supply means.

The drive lines can be conveniently located on either side of the transmission with outputs exiting through the respective primary and secondary drive lines, providing an extremely simple, compact and lightweight powertrain I do.

If the power unit consists of a gasoline or diesel engine, the power transmission unit is conveniently located in the center, directly below or on the same side of the power unit, and the primary and secondary drive lines are, as already mentioned, hollow gears, Input via sprockets or radial or axial hydraulics with hollow center.

For transformer axle front wheel drive applications, the power transmission unit is advantageous in terms of the drive line and power unit located between the front wheel assemblies to accommodate the location and space required for current mass production vehicles. Placed in a good location.

In the case of a four-wheel drive automobile, tractor or truck, the power transmission unit passes through the respective primary and secondary drive lines and through the axially extending holes of the first and second bevel gears of the main transmission. Power output member, such as a half axle, which is conveniently located in relation to aligning the power unit and drive line to align the power output and power transmission in a straight line, thus extending forward and rearward of the differential gear assembly. And then provide rotational power through universal and torque tubes to the front and rear differentials and axles from the front and rear for all four wheels.

In the case of a two-wheel drive truck or a rear-wheel drive truck, the power unit can be left in the vertical position in the usual traditional way, and the primary and secondary drive lines are the normal Powers the power transmission in position directly. This allows heavy load, high torque trucks and tractors that typically require a 10 to 12 speed gearbox to optimize engine speed and torque under the heavy load and high torque required over a wide range of operating conditions. Providing a freely rotatable pinion inside the annular support of the outer transmission member means that simply adding an additional crown gear and pinion to the gearbox can be eliminated from the standard crown gear and pinion.
This provides a continuously variable transmission from reverse to overdrive by simply changing the ratio between the primary and secondary drive lines.

When the power unit is a rotary, tuned rotary, gas turbine or electric engine, the power transmission unit may be configured such that the power unit member passes through the center of the hollow rotor or turbine shaft, or the hollow secondary opposite the transmission unit. As extending through the drive system,
The primary drive line input can also be conveniently mounted in the center between one side power unit and the opposite side secondary drive line as described above. Further, the housing of the power transmission unit can be integrated with the common casing of the power unit and the secondary drive line.

[0035] Preferably, the internal combustion engine can also supply the primary direct drive line and the secondary shift drive line from a common shaft that forms two shift drive lines.

The advantage of this arrangement is that a standard mass production engine can be used as a single power source, or in some applications a flywheel arrangement can be provided as a means to conserve mechanical energy. The flywheel can be associated with the power unit,
Or it can be unrelated. The flywheel arrangement can be used as an input to supplement power during peak demand and / or to provide regenerative braking. The flywheel can also be used as a single power source or directly driven by regenerative braking, using a fixed gearing and overrun sprag clutch between the engine and the flywheel, and the flywheel is then To power the primary and secondary drive lines via Alternatively, the flywheel can be positioned as a high gear position on the primary or secondary drive line and lowered by a suitable gearing to the transmission unit primary or secondary drive input.

As described above, the power transmission unit continuously varies the input rotational power to provide maximum fuel efficiency, minimum pollution and smooth vehicle performance while operating a single power source at constant speed and torque. Combines the functionality of the main transmission and the differential gear assembly conveniently so that the two rotational power outputs can be differentially transmitted.

The control means preferably comprises a microprocessing control unit having an input device for receiving a command input, for example from a vehicle operator, and a plurality of means for providing closed-loop feedback control of the performance parameters of the vehicle powertrain. It consists of an input / output interface device. The plurality of input / output interface devices preferably comprise a plurality of high performance sensors for monitoring, analyzing and transmitting data on powertrain performance parameters. Preferably,
The performance parameters are continuously controlled by the microprocessing control unit and include performance parameters specific to the type of power unit including the variable secondary driveline hydraulics and interaction forces and the controls that make up the vehicle powertrain. For example, if the power unit is comprised of one internal combustion piston engine, the performance parameters continuously controlled by the microprocessing control unit may further include maximum efficiency fuel combustion and minimum pollution, such as manifold pressure and / or boost pressure, Engine torque, engine speed, fuel-air mixture, fuel flow, sprocket timing, valve timing, variable intake manifold shape, variable compression, variable precombustion chamber compression, combustion chamber condition, compression ratio and compression ratio for indirect ignition diesel engines Includes performance parameters specific to exhaust gas chemistry and temperature.

In use, the microprocessing control unit preferably provides a self-diagnostic closed loop feedback control that continuously monitors, analyzes and interdependently adjusts performance parameters in response to command input from a driver. I do. In particular,
The microprocessing control unit continuously controls the power unit speed and torque and other power unit variables, preferably in response to command input and / or analysis of the performance parameter data, to reduce fuel combustion efficiency and minimal pollutants. Maintain, while interdependently adjust the performance parameters including relative speed of power input means and hydraulic pressure and speed of primary drive line and secondary variable drive line and reactive load sharing between transmission and two drive lines and power unit By doing so, the final output speed and power of the power train meeting the operation requirements are continuously controlled.

Preferably, the microprocessing control unit is programmable with a performance algorithm, and thus continuously adjusts the controlled performance parameters according to an algorithm that optimizes powertrain performance. For example, the microprocessing control unit can be programmed to optimize powertrain efficiency. In that case, in response to command input from the vehicle driver, the microprocessing control unit continuously monitors, analyzes, and interdependently adjusts the performance parameters of each of the two power drive lines from the single power unit. According to the load sharing between the power unit and the primary drive line and the reaction load transmitted from the transmission by changing the length of the rod of hydraulic pressure and trapped oil to maintain the efficiency of the power unit within the peak range By simultaneously and continuously monitoring and adjusting the tolerable amount of drive speed reduction of the decreasing secondary drive line, the final output speed and power of the transmission to meet operating requirements will be interdependently controlled.
In such a function, overall power unit efficiency would be achieved over a wide range of different operating conditions.

It will be appreciated that if the power unit is of the internal combustion engine type, a significant improvement in fuel economy and a corresponding reduction in exhaust emissions will be achieved. In addition, to optimize overall power unit efficiency, the power unit can be a rotary engine, a rotary engine adjusted for compression and expansion cycles, a gas turbine engine, a diesel engine, a gasoline engine, an electric motor, or a power unit with an energy storage and regenerative braking system It will be understood that hybrid combinations of can also be included. This makes it possible, for example, for a generator and electric motor power unit with an internal combustion engine and a high-energy and low-weight effective battery storage device, or an internal combustion engine with a combined flywheel regenerative braking power unit, a generator and a combined motor / generator flywheel An integrated internal combustion engine may take the form of energy storage and regenerative braking as one power unit. As a further example,
Preferably, by using a combination of hybrid power unit technologies in the present invention, such as a fuel cell can be used, which converts methanol to hydrogen, which can be supplied to the fuel cell through a proton exchange membrane. The fuel cell, in combination with atmospheric oxygen, provides motor power, which is then fed to the motor / generator flywheel for stored energy and regenerative braking. This power unit supplies power to advanced transmissions as described, provides a means of real-time interactive precision control of power use and regeneration for optimal efficiency, and pollution can be eliminated only by water vapor emissions Will. Alternatively, the fuel cell and regenerative braking provide energy to an effective energy storage battery that provides power to an electric motor that is connected either directly or via an energy storage flywheel to an advanced transmission as described above. Could be supplied. Transmission microprocessor control prevents system overload in response to externally changing load demands, and allows immediate adjustments to derive optimal input power matched to the power unit to provide optimal power output performance. There will be.

In another preferred use of the confinement oil, which is under pressure due to the reaction force of the transmission in the secondary drive line, energy storage and regenerative braking energy may have short peaks such as stop / start in city driving. It can also be stored by oil compressed nitrogen in a nitrogen-filled accumulator that provides the required energy.

In an internal combustion engine, the ignited fuel-air mixture burns in a flame front that extends to the periphery of the confinement area of the cylinder. The generated inflation gas pushes the piston away from the cylinder head, which turns the crankshaft and provides a power stroke. As the piston moves away from the inflation gas, the cylinder volume increases. The higher the initial cylinder compression ratio, the higher the combustion rate and the corresponding expansion rate, requiring a higher piston speed comparable to a higher engine speed. In a fixed compression engine, it is desirable to have a piston speed that matches the rate at which the gas expands at the optimum cylinder pressure, resulting in optimum fuel efficiency and lowest pollution levels.
The aim of the preferred embodiment of the present invention is to utilize control of a variable oil rod that can vary the proportion of transmission reaction force to vary the ratio between primary and secondary drives from a common power source. By doing so, the reaction force is balanced by the initial drive from the common drive shaft, and the speed and torque of maximum efficiency to meet variable load demands and also match the energy levels required to maintain the desired vehicle speed. In order to maintain the optimal balance, the speed fluctuation of the primary and secondary drive lines is applied to the outer transmission, and the input power source is continuously adjusted. As described above, by having the function of a stepless speed constant mesh transmission, even if the engine speed changes and the compression ratio and the fuel-air mixing ratio change, the optimum fuel-air combustion ratio is maintained and required. Providing precise vehicle speed, optimal driving conditions can maintain vehicle speed in real time over the full range of energy demands, providing minimal pollution and maximum fuel efficiency. Further control of the speed of the frame front, in particular,
Related to diesel engines, but having a variable volume pre-combustion chamber and / or main fuel chamber. The interaction of the input and output from the microprocessing allows for optimal setting of engine speed, combustion chamber compression ratio and / or pre-combustion chamber displacement, along with the corresponding mechanical movement and speed of the piston or rotary engine restraint. Fuel air combustion speed (frame front)
Until smooth.

Best Mode for Carrying Out the Invention Embodiments of the invention will be described by way of example only with reference to the accompanying drawings. The principle of operation of the transmission forming part of the present invention will first be described with reference to FIGS.

Differential ratio and reaction force: 3: 1 ratio of input: As can be seen from FIG. 4 by way of illustration, when the rack 1 moves forward by 30 teeth in the direction of arrow A, and when the spindle 4 of the gear Power input ratio 3: B direction:
When moving forward a distance equal to 10 teeth of rack 1 equal to 1, gear 2 causes a rotation of 20 teeth clockwise, so that gear rack 3 moves 20 teeth in the direction of arrow C.
Only the teeth move. On the other hand, as the gear 2 moves forward by a distance of 10 teeth in the direction of arrow B, the rack 3 will move by a distance of 10 teeth in the direction of arrow C. The overall effect is that the primary drive moving rack 1 moves forward and the pivot 4 of the secondary drive moving gear 2 moves forward at a speed of 1/3 at a 3: 1 ratio with the primary drive rack 1. Rack 3 moves in the opposite direction at 1/3 the input speed of primary drive rack 1.

Due to the reaction force associated with the 3: 1 ratio, the rack 1 requires a force in the direction of arrow A, the spindle 4 of the gear 2 must be restrained by the force in the direction of arrow B, and the gear 4 2 and also causes a force on the rack 3, ie a force acts in the direction of arrow C. Both the rack 1 and the gear spindle 4 move in the same direction in a 3: 1 ratio, but restrained the gear spindle 4 in the opposite direction to the arrow B and in the direction opposite to the arrow C. Only when is it clear that the rack 3 acts in the opposite direction.

In FIG. 5, when the rack 1 moves forward by a distance of 30 teeth in the direction of arrow A,
Also, when the gear spindle 4 moves forward in the direction of arrow B by a distance equal to the 15 teeth of the rack 1, this will give a 2: 1 ratio between the primary and secondary power inputs. If the gear spindle 4 has moved half the distance of the rack 1, then the gear rotates 15 teeth clockwise around the spindle 4, but as the gear spindle 4 moves forward a distance equal to 15 teeth, the rack It will be seen that 3 remains stationary at C. It can be seen that the gear rack 3 remains stationary in the dynamic neutral lock position by moving both the primary drive rack 1 and the secondary drive line gear spindle 4 in the same direction at a 2: 1 ratio.

The reaction force for the 2: 1 ratio between the primary drive rack 1 and the secondary drive line to the gear spindle 4 has no necessary reaction force other than the friction that occurs.
No load is applied to any direction of the output rack 3 which is stationary at C.

If a load is applied to the rack 3 in the direction of the arrow B, a force with a reaction force equal to the gear spindle B in the opposite direction to B will be required to balance the force on C, When a load is applied to the rack 1 in the direction of arrow A, the reverse rotational force transmitted from the force acting on the rack 3 at C and acting on the gear 2 around the spindle 4 which is equal to the force acting on the rack 3 must be equal. There will be. The overall effect is that the forces required on the gear rack 1 in the direction of arrow A need to be equal and on the gear rack 3 via the gear 2 rotating in rotation about the spindle 4 on the gear rack 3 in the direction of arrow B. A force equal to the force will be transmitted, so that the spindle 4 will receive the sum of both forces on the racks 1 and 3 and will need to be restrained in the opposite direction to the arrow B. This means that the restraining force required for the spindle B is doubled with respect to the load applied to the output rack 3 at C.

When a reverse load is applied to the power output gear rack 3 at C in the opposite direction of the arrow B,
In this case, a double force is required for the spindle 4 in the direction of arrow B, and a reaction force is applied to the rack 1 in the direction opposite to the direction of arrow A.

Due to the 2: 1 ratio of the primary and secondary power drive lines, the rack 3 remains stationary but depends on the direction of the load on the gear rack 3,
Even if both the primary and secondary power drive lines turn in the same direction and the reaction energy of the load on the output rack 3 is reversed, the direct reaction force in opposite directions on the primary drive line rack 1 and the secondary drive line gear spindle 4 As a result, the load that turns off the forward driving restraint is reversed, and the direction of the load on the power output rack 3 at C alternately changes.

In FIG. 6, when the primary power drive line input rack 1 moves forward by 30 teeth in the direction of arrow A, and the secondary power drive line moves the spindle 4 in the direction of arrow B in FIG. When moved, the rotation is constrained by the same distance that the sprocket 2 is moved by the rack 1, and the gear spindle 4 and the gear 2 and the teeth of the gear 2 mesh with the teeth of the rack 1 to prevent rotation. As a result, the gear rack 1
And the sprocket 2 both move in the directions of arrows A and B at the same speed as a fixed mass. The gear 2 moves at the same speed in the direction of arrow C while the teeth of the gear 2 mesh with the rack 3.

In most vehicle operations, the ratio is 1: 1 and the transmission only intermittently receives reverse thrust in the direction of arrow C of rack 3 in FIG. In such a case, the vehicle has been decelerated using compression of the engine as a brake or by means of an exhaust brake. In all other cases, the rack 3 will move in the opposite direction to arrow C and will experience a force of varying magnitude in the opposite direction. This force is restrained by the meshing of the teeth of the gear 2 meshing with the teeth of the rack 3. The force exerted on the gear 2 by the rack 3 in the direction opposite to the arrow C causes the gear 2 to rotate clockwise around the spindle 4, which is in a direction opposite to the arrow A in a 1: 1 ratio. It will be restrained by the gear rack 1 with the required restraining force. The force required on the spindle 4 in the direction of arrow B is opposite to the movement of the rack which is opposite to arrow C, in the opposite direction to the vehicle driving load and the direction of movement of the rack 3 and to arrow A. The restraining force on the rack 1 in the opposite direction will be the sum of the force required to restrain the gear 2 from rotating about the spindle 4.

The force required for the secondary drive line at spindle 1 at a ratio of 1: 1 is indicated by arrow B
And the force on the primary drive line 1 will be equal to twice the force acting on the power output rack C in the opposite direction to arrow C, and
And A will be equal in the same direction as the force acting on the power output rack 3 in the opposite direction.

1: 1.5 Ratio All ratios above the 1: 1 ratio can be selected as overdrive, but have also been selected for purposes of illustration of the 1: 1.5 ratio. Using this ratio as shown in FIG. 6, when the primary drive line rack 1 moves forward by a distance of 30 teeth in the direction of arrow A, and the spindle 4 moves 1.5
As the spindle 4 of gear 2 moves 15 teeth further than rack 1 when gear 4 moves forward a distance equal to a total of 45 teeth, a total of 45 teeth, gear 2 moves 1 counterclockwise.
It turns only 5 teeth. The combined acting force is the same 30 tooth distance that the power output rack 3 is moved by the primary drive rack 1 plus an additional 30 tooth distance caused by the sprocket 2 counterclockwise direction. Over a total distance of 60 teeth, the rack 3 will move in the direction of arrow C, and will therefore cause overdrive twice that of the primary drive line speed.

FIG. 7 for illustration purposes shows the primary power input drive line as a straight rack 1 and the primary drive line as A. The secondary power input drive line is shown as spindle 4 and gear 2, which can freely rotate around spindle 4 and is denoted as secondary drive line B. The combined result of the combination of the ratio and the reaction force has already been described with reference to FIGS. 4, 5, and 6, and the resulting combined power output is shown as C on rack 3 in FIG. The flat linear form of the racks and gears is used for illustration purposes, where racks 1 and 3 have been replaced by crown gears 5 and 6 and gears and spindle 4 have been replaced by pinions 7 and bearings 8. It will be readily appreciated that the same ratios and reaction forces will still be applied.

As the rack 1 moves forward in the direction of arrow A and the crown gear 5 moves in the same direction as the rack as viewed from the top surface, the crown gear rotates counterclockwise on the crown gear 5 in the direction of the arrow. The pinion 7 has the same effect as the flat profile gear 2 and turns clockwise when viewed from the top, and the pinion 7 turns the crown gear 6 at C in the direction of the arrow of the power output crown gear.

As described in FIGS. 4, 5 and 6, the secondary power input drive line moves the spindle 4 and thus the gear 2 in the direction of arrow B as shown in FIG. The same effect is obtained by turning a pinion 7 complete with a bearing assembly 8 about the central axis of the crown gears 5 and 6 turning about the axes AC.

For ratios and reaction forces, for ratios of 3: 1, 2: 1, 1: 1 and 1: 1.5
However, the output speed fluctuation and the reaction force are calculated according to the following formula.
It is understood that it is proportional to. V_out= 2 × V_secondary-V_primary  Where V_outIs the output speed, V_secondaryIs the secondary input
Speed (planetary gear assembly) and V_primaryIs the speed of the primary input
Yes, with outputs corresponding to the vector analysis of the reaction forces of the two input means.

The present invention provides an infinitely variable transmission from reverse to full speed through stationary dynamic lock and overdrive. The generated reaction force is A
Must be constrained between the primary power input drive line represented by and the secondary power input drive line represented by B, so that they react against the common drive shaft, thus avoiding energy loss. ing. To achieve this, the present invention provides a differential variation in the amount of protrusion of the trapped oil rod against speed variations between the power input drive line A and drive line B or any such suitable The pinion 7 allows fluid to be controlled and imparts the opposite differential hydraulic action to freely rotate the bearing 8 and, as shown in FIG. 7, the crown gears 5 and 6 which rotate about the axes AC. It provides shifting and control of the reaction force generated by the rotation of the surrounding pinion bearing assembly 8.

This is shown in more detail in FIG. The primary power input drive line is
It is operatively connected to a bevel gear 5 indicated by arrow A. Secondary power input drive
Line is controlled by differential variable speed means and variable reaction torque is confined
Oil by dynamic differential reverse hydraulic control or appropriate fluid
Controlled and rotatable or constrained to the bearing 8 depending on the reaction force
, The pinion gear 7 in the bearing 8 rotates as B around the axis AA
It rotates freely as the reaction opposing rotational force expressed. From a further viewpoint, the present invention
In connection therewith, the differential 9 disclosed in the applicant's earlier application
It is included in the inside of the package 10. Output speed and force are input means according to the following formula
Corresponding to the speed. V_out= 2 × V_secondary-V_primary  Where V_outIs the output speed, V_secondaryIs the secondary input
Speed (planetary gear assembly) and V_primaryIs the speed of the primary input
Yes, reaction force vector of primary drive line A and secondary drive line B
The output transmitted to the bevel gear 6 for the output corresponding to the analysis is expressed as output power C.
It is shown. The bevel gear 6 transmits the rotational force to the central differential by the rotary carrier 20.
And represented as D.

The present invention is a proven and advantageous combination of crown gears, pinions and differential components that have been demonstrated and used in a wide range of vehicles, trucks and tractors, but with reduced speed and reaction forces. It uses a unique and novel combination of controls.

FIG. 9 shows a power transmission 10 according to a preferred embodiment of the present aspect of the present invention, which generally has two rotatable input means operatively connected to rotatable output means. It comprises a main transmission 11 and an internal differential gear assembly 9 having rotatable input means operatively connected to two differential rotatable output means. These components will be described in further detail with descriptions of the various preferred embodiments.

The main transmission 11 of the power transmission 10 preferably comprises a gear train. The two input means of the main transmission 11 are the first bevel gear 5
And a planetary gear assembly 1 coaxially arranged to rotate about a first axis AA
Consists of two. The planetary gear assembly 12 comprises an annular pinion carrier 13 and a planetary bevel pinion gear 7 disposed inside which the annular pinion carrier 13 rotatably carries, the planetary bevel pinion gear 7 being perpendicular to the first axis A-A. Has an axis. Ninth
In the figure, as indicated by the letters A and B, the first bevel gear 5 and the annular pinion carrier 13 are independently provided by conventional means such as gears, belts, chains or directly to the power input primary and secondary drive lines, respectively. Operablely connectable. The output means of the main transmission 11 comprises a second bevel gear 14 arranged coaxially with both the first bevel gear 5 and the planetary gear assembly 12, and the first shaft A-
Rotate around A. First and second bevel gears 5 of main transmission 11
, 14 each have a centrally formed hole extending in the axial direction. The planetary gear assembly 12 is disposed between the first and second bevel gears 5, 14, and includes the first and second bevel gears.
With a planetary bevel pinion gear 7 meshing with both bevel gears 5,14.

The differential gear assembly 9 is arranged inside the main transmission 11. The input means of the differential gear assembly 9 of the power transmission 10 consists of a differential bevel pinion gear 15 arranged radially inside the main transmission 11, about an axis perpendicular to the first axis A-A. To rotate. The output means of the differential gear assembly 9 comprises two differential bevel side gears 16, 17 coaxially arranged to mesh with the differential bevel pinion gear 15 and rotates about a first axis A-A. . The two differential bevel side gears 16, 17 are mounted in the center of opposite ends of power output members 18, 19, such as two coaxially aligned half axles, which are the first of the main transmission 11. It extends axially outward through holes in the first and second bevel gears 5,14. The half axles 18, 19 operatively connected to the drive wheels (not shown) or, in the case of four-wheel drive, to the power transmission unit, can also be mounted centrally, in which case the half axles 18 , 19 are operatively connected to the front and rear differentials. The main transmission 11 is operatively connected to the differential gear assembly 9 by a differential frame 20, which is connected to a second bevel gear 14 of the main transmission 11, which The moving pinion gear 15 is carried.

FIG. 10 shows a practical embodiment of a power split device according to the present invention used to provide power input primary and secondary drive lines from a single power source to a transmission. This device provides variable speed differential displacement and dynamic mechanical / hydraulic control as well as control of the reaction force between the power input primary and secondary drive lines. The reaction force, which is controlled to avoid power loss by reacting against a common shaft having a single power input, is directly connected to a single power source or a gear, V-belt, multi-V power belt (With a common outer flat belt), toothed belts, rollers or connected by conventional means such as by special high speed or silent chains or hydraulic output.

FIG. 10 is for illustrative purposes, and in one preferred embodiment, a power unit such as a gasoline internal combustion engine or a diesel engine is operably connected by a gear 20 to turn a balanced assembly 21 and 20 is fixed to a housing 22, which transmits a primary input power line to a gear 23 via a rotatable housing. The gear 23 is fixed to the housing. The housing is carried by bearings 24 on a fixed shaft 25 whose rotation is restricted by fixtures 26. The housing is further carried at opposite ends by bearings 27. By this means, power is transmitted from the power unit through the primary drive line, through the gear 20 through the housing 22 and the gear 23 at point A.
Directly transmitted to Output speed depends on engine or power unit and assembly 2
The speed change of the input power unit is directly determined by A
Determined by point gear 20.

A second shaft 28, which is separately rotatable and independent of the shaft 25, is carried by a bearing 29 and has a hydraulic / mechanical differential through a gear 30 providing a power input secondary drive line at point B. It is directly connected to the power transmitting gear 30.

The fixed shaft 25 has a stationary circumferential piston group 31, and the rotation of the stationary circumferential piston group 31 is restricted by splines of the fixed shaft 25. A rotatable shaft 28 carried by a bearing 29 is splined to the rotatable piston group 32 and is thus operatively connected to the piston group 32 with the shaft 28 and a secondary driveline gear or sprocket 30 and at point B Provides reaction force. The rotatable housing 22 itself has a common porting and commutator plate 33 mounted between the piston groups 31,32. The common porting and commutator plate assembly 33 is fixed to or rotatable with the housing 22. This ensures that high and low pressure kidney bean-shaped porting is always present on each side of the circumference of the assembly 33 and that both the reaction displacement rotating groups 31, the swash plate control swash plate 35 and the rotating housing 22 are associated.
And the fixed pivot point 3 of the fixed displacement swash plate 36
Align with 4.

The present invention provides differential hydraulic / mechanical variable speed and torque reaction control between the primary power input fixed drive line A and the secondary variable power drive line B. This is accomplished by a primary driveline input gear or sprocket 20 at input A fixed to housing 22 and rotatably and operably connected at A directly to primary driveline sprocket or gear 23 for primary power supply. The torque force and rotation speed, which are directly controlled by the speed and torque of the input power unit to the gear or sprocket 20 at A, are transmitted directly to the gear or sprocket 23 at A. The reaction force transmitted from the secondary power input B acts directly on the sprocket or gear 30 at B as shown in FIGS. 4, 5, 6, 7 and 8 as shown in FIG. The gear 30 is restricted in rotation by a shaft 28 and a rotating group 32, which in turn
Rotation is constrained by the control of confined oil in a cylinder 37 displaced by a piston 38 that contacts the fixed angle swash plate 36 as the reaction force rotates the piston group 32 in relation to the fixed angle swash plate 36 and the oil 37 The confinement rod is displaced by the piston 38 as the piston rotates toward the cylinder position 35. The allowable amount of rotation is determined by the variable angle of the swash plate 35, the displaced oil moves freely through the kidney-shaped porting on each side of the circumference of the commutator assembly 33, and the commutator assembly 33 remains fixed. Until,
Rotates with housing 22 corresponding to swash plate pivot point 34. The oil rod confined in the cylinder 37 is almost 180 ° around the circumference.
Oil can be transported freely through the elongated bean-shaped portion and as the rotating piston assembly 31 is stationary and the swash plate 35 is driven by the rotation of the housing 22, oil is transferred through the commutator assembly 33 to the cylinders 40 and 41. It is understood that it is transferred to. As the swash plate moves to cylinder position 41, the displacement space of cylinder 40 increases,
Therefore, it is possible to control the transfer of the oil trapped in the expansion cylinder between the cylinders 37 to 31, 41. When the primary input gear 20 turns the housing 22,
Next, the swash plate 35 is driven through the passage to the fixed swash plate pivot point 34 in response to the housing 22 and if the swash plate control angle is set to position 1 equal to the fixed angle of the swash plate 36, the rotation The displacement of the rod of confined oil in the piston assembly 32 will be equal to the expanding displacement chamber of the fixed rotating piston assembly 31, so the reaction force from the outer transmission acts on the gear 30 at B and the shaft It will be seen that it rotates rearward via 28 in response to primary sprocket 20 and housing 22. This is gear A at point A
Under a full speed input to zero and a full power input, the output speed of the gear 30 will be zero and the reaction force will be a direct hydraulic / It means that the mechanical variable control is completely restrained without energy loss.

FIG. 11 shows the same general arrangement as that of FIG. 10, but with a swash plate 35 controlled around the axis 34 by adjustable means in the operating position 4. As power is applied at the input primary drive line A, the gear 20 and housing 22 rotate, which also actuates the swashplate 35 at the same ratio of 1: 1 regardless of the angular position of the swashplate 35. If the angle of the swash plate 35 is set to 90 ° with respect to the center line of the axis 25 at the position 4, then there is no change in displacement between the piston chambers 40, 41 and the displacement from the cylinders 37, 39 All flows of confined oil will be zero. As power is input to the primary drive line, A
Through the housing 22 and the gear 23 to the input 1 in the outer mechanical transmission primary drive line as shown in FIGS.
The reaction force described in FIGS.
Acts on zero. The gear 30 is constrained by the shaft 28 and the zero displacement change in the fixed piston assembly 31 prevents the rotating piston group 32 from transferring the rod of confined oil in the syringe 37 with the piston 38 through the commutator assembly 33. The combined effect is that the entire assembly 21 rotates together with the primary power input and output sprockets 20, 23, and the secondary reaction power control sprocket 30 rotates in B all in a 1: 1 ratio.

FIG. 12 is also similar to FIGS. 10 and 11, but further shows the five operative positions of the swashplate 35 for illustrative purposes. In position 1, the output gear or sprocket 23 supplies power to the A outer transmission sprocket 42 via the primary drive line in direct correlation with the speed and torque provided by the power unit as described in FIGS. I do. If the control angle of the swash plate 35 is located at the control position 1,
Next, the function described with reference to FIG. 10 occurs, and the reaction force is transmitted at B through the annular pinion carrier 13 and connected to the sprocket or the gear 43.
It controls the variation of input and output loads and the reaction force, which varies directly with the direction of the force, while still maintaining zero rotation as described in FIGS. 4, 5, 6 and 7.

When the control angle of the swash plate 35 changes to the position 2, the reaction force applied to the gear or the sprocket 30 by the displacement is controlled by the secondary power drive line at the point B, and the reaction force on the primary power drive line at the point A / 3
become. This produces a 3: 1 ratio, which is input into the main transmission 11 and is transmitted at a speed and energy transmitted to the bevel gear 14 of C via the bevel pinion gear 7 and at a reverse output via a transfer of energy, 1 / of the input speed. Give 3.

FIG. 12 sets the washplate angle at position 3 and shows that the displacement rotates the secondary power drive line at B at half the speed of the primary power drive line A at B. The effect of this reaction force enters the main transmission 11 at a 2: 1 ratio as described in FIGS. 4, 5, 6 and 7 and creates a dynamic lock neutral position with zero output from the main transmission 11 and a twelfth. The swashplate 35 is set at position 4 as shown, indicating that it is at 90 ° to axis 25. The effect of the reaction is input between the primary power drive line A and the secondary reaction power control drive line B at a 1: 1 ratio, as fully described in FIG. Therefore, one of the inputs
The 1: 1 ratio controls the correlation between the speed of the main transmission and the torque force interacting with it, producing a 1: 1 full forward output from the transmission via the second bevel gear 14 of C, and the combined effect is shown in FIG. The oil transfer flow from the piston rotation group shown in FIG. There is no movement of the piston rotation group 32 relative to the housing, all components in the working displacement hydraulic / mechanical reaction power control unit 21 rotate in unison except for the fixed shaft 25, and the fixed piston group 31 is always stationary. It remains.

By supplying an equal speed of 1: 1 on the primary drive lines A and B and controlling the reaction energy by the power control drive line B, all components of the main transmission 11 have zero movement. Thus, for vehicle, truck or tractor applications, there is only a 2: 1 variation at zero speed, and the primary and secondary power drive lines have a speed with the secondary drive line B that is half the speed of the primary power drive line. And 1: 1
The instantaneous and accurate adaptation to the variable load can be achieved with a smooth transition to a ratio of.

As shown in FIG. 12, when the control position of the swash plate 35 has advanced to the position 5, a ratio of 1: 1.5 can be executed next, and as a result, the overdrive output from the C main transmission Results in a primary power input drive line rotation speed equal to 1.5 times, as already described in FIG.

A further aspect of the present invention, shown in FIG. 12, is that centrifugal force causes the pressure of the oil and the confined oil mass toward the outer periphery of the rotating housing 22 as the rotating housing 22 filled with hydraulic fluid rotates. Is to create it. This pressure, in one embodiment of the invention, acts as a filling pressure to fill the radially located cylinders 40, 41, 37, 39 and the radially located cylinders therebetween, and Are also arranged in the zone of high-pressure hydraulic fluid on the outer circumference of the piston groups 31, 32. The normal use of high and low pressure check valves can be used, and can also be used in certain applications of the integrated pump.

In a further embodiment (not shown) of the present invention, the rotating housing 22 has extended cooling fins and / or rotary seals to allow hydraulic fluid to enter and exit the housing 22 for filtration and cooling. In preparation.

In another embodiment, the rotating piston group and / or the commutator plate will be made of a ceramic material for high speed, high pressure, low lubricity and high effective performance.

The variable angle swashplate 35 can be controlled by a suitable arm and thrust bearing or a suitable perforated pressure control device or a remote wireless or infrared control device to control the angle of the swashplate inside the rotating housing 22. .

FIG. 13 shows an actual embodiment of the arrangement of FIGS.
Includes the mounting configuration required to operate the seal and assembly. These components may in fact be modified, and a description is not necessary to understand the present invention, even if its function would be readily apparent to those skilled in the art. Will.

In use, the main transmission 11 described above comprises a continuously variable transmission, wherein the output speed (the rotational speed of the second bevel gear 14) is the sum or difference of the two input speeds supplied from the integrated assembly 21. A and B are proportional to
(Controlling the rotation speed of the first bevel gear 5 and the planetary gear assembly 12). Therefore, if the speeds of the first bevel gear 5 and the planetary gear assembly 12 are controlled, the speed of the second bevel gear 14 will pass from design maximum through zero in the first direction of rotation and in the second opposite direction of rotation. It will be appreciated that it can be varied continuously up to the design maximum. Also, as described above, the differential gear assembly 9 operates in the manner of a conventional automotive differential gear. The power transmission unit 10 combines the main transmission 11 and the differential gear assembly 9 in a functionally convenient manner as described above, so that the input rotational power is continuously variable output speed to the two half axles 18,19. It can also be transmitted differentially.

The operation of the power transmission unit 10 described above can be further understood with reference to an exemplary embodiment in FIG. 9, wherein the first bevel gear 5 and the second bevel gear 5 of the main transmission 11 14 have 30 teeth each, and the planetary pinion gear 7 of the main transmission 11 also has 10 teeth each.

The effect of the ratio of the input difference in the same rotational direction about the axis AA of the exemplary embodiment illustrated in FIG. 9 will now be described.

A 1: 1 ratio of the inputs from the primary and secondary power drive lines A and B connected together to form a hydraulic / mechanical reaction power control unit 20 has the first bevel gear 5 shown in FIG. Makes one rotation, the 30 teeth of the bevel gear 5 rotate forward. At the same time, when the annular pinion carrier 13 rotates around the A-A axis in the same direction one rotation forward, each planetary pinion gear 7 remains stationary. As a result, the non-rotating bevel pinion gear 7 meshed with the second bevel gear 20 completely turns the bevel gear 14 and the differential frame 20 forward around the A-A axis by one turn. The differential pinion gear 15 carried on the differential frame 20 is restrained only by differential side gears 16, 17 operatively connected to half axles 18, 19, respectively, whereby two drive wheels (shown). ) Can perform the functions of a conventional motor vehicle differential.

The results shown in FIG. 13 indicate that the ratio between the input provided by the differential hydraulic / mechanical reaction power control unit 21 shown in FIG. 13 between the first bevel gear 5 and the annular pinion carrier 13 The speed ratio through the power transmission unit 10 to the differential frame 20 is 1: 1.

When the first bevel gear 5 shown in FIG. 9 makes one forward rotation, the input ratio of the differential power control unit 21 and the power transmission unit 10 is 2: 1. A half turn (2: 1) forward around the AA axis in the direction:
Each planetary umbrella pinion 7 rotates 15 teeth around the axis perpendicular to the AA axis in the reverse direction,
From the point at which it meshes with the second bevel gear 14, it moves by half a rotation of 15 teeth. The second bevel gear 14 is 3
Since it has zero teeth, the annular pinion carrier 13 and the bevel gear 7 move half a turn, so that the second bevel gear 14 with 30 teeth and the differential frame 20 remain stationary.

As a result, as shown in FIG. 13, the ratio of the input provided by the power control unit 21 is 2: 1 between the first bevel gear 5 and the annular pinion carrier 13 and the second bevel gear 14 And differential side gears 16 and 17 remain in an effectively locked rest position with zero output from power transmission unit 10.

When the first bevel gear 5 makes one full rotation for 30 teeth forward, the annular pinion carrier 13
Rotate one-third forward around the A-A axis in the same direction, each planetary bevel pinion gear 7 reversely rotates backward by 20 teeth, during which a distance equal to ten teeth around the second bevel gear 14. Move. For this reason, the second bevel gear 14 and the differential frame 20 have 10 teeth in opposite directions,
Or turn 1/3 in the opposite direction. As described above, the differential gear assembly 9 is capable of performing normal motor vehicle differential action between two drive wheels (not shown).

As a result, the input ratio between the first bevel gear 5 and the annular pinion carrier 13 is 3: 1.
And the second bevel gear 14 and the differential frame 20 are １ ／ in the opposite direction around the AA axis.
Rotate.

It will be appreciated that overdrive is obtained with an input ratio of less than one. In view of the above, the implementation of the integrated power transmission unit of the present invention provides a compact, integrated, continuously variable transmission with a conventional motor vehicle differential, which is particularly, but not exclusively, also for front-wheel drive. Mid-mount for vehicles, four-wheel drive vehicles, vehicle power trains for trucks and tractors, replacement of rear-wheel drive differential assembly according to the present invention eliminates the need for a gearbox, and the present invention provides vehicle power for vehicles, trucks and tractors. It will be appreciated that the train is suitable for use. In particular, embodiments of the transmission of the present invention allow two vehicle drive wheels to be continuously rotated at variable speeds for differential drive or front and rear axles to be differentially driven.

As shown in the accompanying drawings FIGS. 8, 9 and 13, generally a differential hydraulic / mechanical reaction power control unit 21 and a power transmission unit 10 according to a preferred embodiment of the second aspect of the invention And a single power input unit integrated assembly connected thereto, and these integrated assemblies have been described above.

FIG. 14 shows the power unit 44, the differential displacement hydraulic / mechanical variable speed torque reaction secondary power drive line unit 21, the transmission unit 10, and the buy close processing control unit 5. These components are now described in further detail for various preferred embodiments.

The foregoing detailed description of the structure, arrangement and operation of the various components of the power transmission unit 10 will be referred to in the following description.

In the embodiment shown in FIG. 14, the power unit 44 is an internal combustion engine and alternatively all types of internal combustion engines—gasoline engines, diesel engines, rotary engines, true rotary balanced engines gas turbine engines ,
Or consist of any similar or dissimilar combination that can also use conventional electric motor locks, fuel cell powered electric motors, electric and / or mechanical flywheels. Gasoline and diesel internal combustion engines are preferred because of their established technology at relatively low production costs. In the embodiment shown in FIG. 14, the power unit 44 is a gasoline or diesel engine power unit.

The hydraulic / mechanical power control unit 21 communicates from the crankshaft at the rear of the engine via gears, chains or belts that fit gears or sprockets 20, and connects the primary and secondary power input drive lines to the engine / Advantageously arranged for transmission from the primary and secondary drives A, B of the power control unit 21 to the drive gears or sprockets 42, 43 of the power transmission unit 10 via gears or chains or belts to the center of the power unit 44, The power transmission unit 10 is centrally located with respect to the horizontal engine 44, which is the most suitable location for front wheel drive applications, and connects the half axles 18, 19 from the standard front wheel drive shaft to the front wheels. Will be placed at the appropriate location.

The housings of the power control unit 21 and the power transmission unit 10 are also advantageously arranged so that they can be easily removed from the engine 44.

FIG. 15 shows a simplified embodiment of the vehicle power train 46 shown in FIGS. 9 and 10, in which the second bevel gear 1 of the power transmission unit 10 is shown.
4 is not operatively connected to the internal differential gear assembly, ie, only the power transmission unit 10 includes the main transmission 11,
However, it is instead operatively connected to the drive shaft 47. A simplified embodiment of the vehicle powertrain 46 shown in FIG.
Utilizing a rotary, true unbalanced rotary or gas turbine engine, adapted for use in front and / or rear wheel drive vehicles, with drive shaft 47 operatively connected to a single conventional automotive differential gear assembly. Have been. Alternatively, the vehicle powertrain 47 shown in FIG. 15 can be a mid-mounted four-wheel drive vehicle, with the drive shaft 47 being a conventional front and rear four-wheel drive differential gear assembly. Operatively connected to the

The microprocessing control unit 45 shown in FIG. 14 includes an input device for receiving a command input from a vehicle driver and a plurality of input / output interfaces for providing closed-loop feedback control of performance parameters of the vehicle powertrain. Device (not shown). Preferably, the plurality of input / output interface devices (not shown) are comprised of a plurality of high performance sensors for monitoring, analyzing and transmitting powertrain performance parameter data. The performance parameters are preferably the load of the engine 44, the load of the power control unit including the rotational speed, the hydraulic pressure and the rotational speed of each of the two primary and secondary power output gears or sprockets 42, 43, the two half axles 18, 19 And the rotational speed of each of the two half axles 18, 19. Preferably, the performance parameters are continuously controlled by the microprocessing control unit 45 and further include performance parameters specific or specific to the type of power unit making up the vehicle power train.

In use, the microprocessing control unit 45 preferably includes a closed loop feedback control for continuously monitoring, analyzing and mutually adjusting performance parameters in response to command input from the driver. During startup of the engine 44, the microprocessing control unit 45 conveniently controls the angular alignment of the disengagement safety clutch and / or swashplate 35 shown in FIG. At a 2: 1 ratio to the drive line, the output bevel gear 14 of the main transmission 11 locks to the stationary dynamic lock position at start-up or at a parking safety position.

That is, the starting microprocessing control unit 45 starts the first control unit 21 only when the rotation speed ratio between the power output gears 42 and 43 is 2: 1.
The corresponding power output shaft of the engine 44 is connected to the gear 20.

From the above description of the operation of the power transmission units 21 and 10, each differential side gear 16 and 17 of the power transmission unit 10
Will therefore be reliably locked in a safe rest position at zero speed. It will be apparent that in most applications a manual override clutch is provided to increase the safety level. From the above description of the operation of the power transmission units 21 and 10, following, for example, in response to a command input from a vehicle driver, the micro-processing control unit 45 adjusts the engine speed to adjust the relative speed of the power output gears 23 and 30. 44 and the operation of the power control unit 21 are conveniently controlled, whereby the differential side gears 5 and 14 of the power transmission unit 10 and thus the two half axles 18 and 19 and the two drive wheels (shown in FIG. It will also be understood that (not shown) rotates in the required direction at the required speed. During operation of the vehicle powertrains 21 and 10, the microprocessing control unit 45 responds adaptively to command input and / or analysis of performance parameter data, and to the relative speed and A of the two primary and secondary power drive lines A and B. And two drive lines 23 and 3 at B
The final output speed and power of the vehicle powertrain 48 meeting the operational requirements is continuously controlled by interactively adjusting performance parameters, including reaction load control between zero.

Preferably, the microprocessing control unit 45 is programmable by a performance algorithm so that the performance parameters controlled according to the algorithm that optimizes the performance of the vehicle power train 48 are continuously and interactively controlled. adjust. For example, the microprocessing control unit 45 can be programmed to efficiently optimize the vehicle powertrain 48, in which case:
In response to command input, the microprocessing control unit 45 continuously monitors, analyzes and interactively adjusts the performance parameters of each primary and secondary power drive line to maintain the efficiency of the engine 44 within peak values. On the other hand, a power control unit 21 that interactively controls the final output speed and power of the power transmission unit 10 to suit the operation supply.
Will instantaneously and continuously monitor and adjust the load sharing between the primary and secondary drive lines. In such a function, the total engine efficiency will be achieved over a wide range of different operating conditions. Thus, it will be appreciated that in embodiments applied to internal combustion engines, a significant improvement in fuel economy and a corresponding reduction in exhaust emissions can also be achieved.

For example, when the power unit is an internal combustion piston engine (as in the embodiment shown in FIG. 14), the performance parameters are continuously controlled by a microprocessing control unit, and are preferably further specific to the internal combustion piston engine. Including manifold boost pressure, engine torque, engine speed, fuel mixture, spark timing, valve timing, variable intake manifold shape, combustion chamber conditions, compression ratio and exhaust gas chemistry.

By providing the function of a continuously variable transmission, it provides the required precise vehicle speed, while. By changing the engine speed and the relative primary and secondary power output speeds of the power control unit 21,
And maintaining the optimum fuel-air combustion rate by changing the fuel-air mixing ratio,
Optimal operating conditions can be maintained in real time over the full range of speed and energy requirements for vehicles, trucks and tractors, providing minimal pollution and maximum fuel efficiency.

In view of the above, the vehicle powertrain embodiment of the present invention provides a compact, adaptively controlled vehicle powertrain, especially for applications in small and medium front wheel drive vehicles where weight and size are critical. It will be appreciated that it is suitable.

FIG. 16 is an embodiment of the same principle as described in FIG. 9, and relates to the combination of the two input means of the main transmission 11 and the differential gear assembly 9 to form the power transmission unit 10. A speed change is required between the bevel gear 5 and the annular pinion carrier 13 to which power is directly input from the primary power input drive line, and the same reaction force is generated and controlled as described in FIGS. 4, 5, 6 and 7.

FIG. 16 shows a compact and integrated combination of the power control unit and the power transmission unit 10 of the present invention, which controls both speed change and reaction force by differential displacement as described above. I do. The cam and rack 49 (a) are fixed to the primary power input gear sprocket or pulley 42, the annular cam and rack are fixed to the commutator plate 50, which has two kidney-shaped commutator slots 51 and 52. ,
This pressure extends to almost either 180 ° on either side of the circular commutator plate and separates well between opposing semicircular commutator slots, preventing hydraulic fluid from moving from the high pressure side to the low pressure side. 4, 5, 6 and 7 are the alternating forces described in FIGS. Each commutator slot 51 and 52 shown in FIG. 16 is permanently fixed in alignment with the cam plates 49 (a), 49 (b) so that the highest point of the circular cam track 49 (b) is two commutators. This is a division point between the semicircular slots 51 and 52. By this means, regardless of the rotation of the primary input sprocket or gear, the cam track 4
9 (a), 49 (b) are secured to the commutator valve plate 50, and the commutator slots 51 and 52 also rotate with the cam track to ensure permanent alignment of the cam track with the commutator plate, and provide hydraulic fluid or oil. Is trapped behind the piston 54 of the cylinder 55 by this means to receive and release hydraulic fluid as the piston rotating group 54 rotates differentially. The reaction force on the cam tracks 49 (a), 49 (b) causes the piston rotating group to rotate in line with the primary power input gear or sprocket 42. The piston rotating group is mounted on an annular pinion carrier 13, which rotatably carries a planetary umbrella pinion gear 7 disposed therein. This produces a change in output speed and a torque force, which is constrained by the rotating piston group 53 in the manner described above.

Separating the high pressure commutator slot from the low pressure side to control two semicircular commutator slots, either a single slot or a series of circular holes in a curved path interconnecting the narrow center slot Since the commutator slots 51 and 52 rotate differentially, it is necessary to offset the stationary commutator gallery 56 at the port A and the gallery 57 at the port B with respect to the A-A axis. By this means, the piston rotating group cylinder oil galleries 51 and 52, which rotate differentially via the commutator plate slots 51 and 52, are always adjacent to the corresponding high and low pressure ports, respectively.

As described in FIG. 16, an advantage of the present invention is that the control of the required change in differential power input speed between the primary and secondary power input drive lines and the control of the interaction force are all Compact integrated unit suitable for retrofitting the differential to the current position of the vehicle, truck and tractor or any such machine that normally requires forward and reverse, shifting and differential action and further removal of the gearbox Has been completed.

As described in FIG. 16, the completely integrated unit is the differential reaction hydraulic / mechanical power transmission 58 shown in this embodiment, and is a reaction mechanism for controlling the operation force by the secondary power input drive line. Is a piston with a ball running on a circular concave cam track. It can also be fitted with a piston ring and rollers attached to the end of the piston by bearings and shafts, the rollers being able to freely rotate around a plurality of rollers and hitting a suitably shaped cam track. Other embodiments using a contactable piston can also be used. The assembly can also be fitted with side rollers to prevent high torque from being transmitted to the piston. One embodiment is shown later in FIG. 23, and a further embodiment is shown in FIG.

As described with reference to FIG. 16, the present invention includes an outer hydraulic control differential power control unit including the mid / main power transmission unit 11 and the internal drive differential device 9.

It will be further appreciated that differential carriers applied to vehicles rotate at a ratio of about 4: 1 with respect to normal engine speed. For this reason, in the differential reaction hydraulic / mechanical power transmission 58 used in the present invention, the so-called 4,000 rpm full speed engine speed only rotates at 1,000 rpm for long life and durability of all components. I do.

The installation of the rotating piston group 53 on the outer periphery ensures high torque at low speeds and the width between the half axles 18, 19 keeps the mounting between the front wheels of the front wheel drive vehicle to a minimum. Or to replace standard differentials for vehicles, trucks and tractors in standard spaces and standard axles without half axles, and to control microprocessor controllers as described in stepless reaction transmissions. Including integrated installation.

In FIG. 16, it is assumed that the speed and the reaction force are controlled by the hydraulic oil pressure and the flow rate of the port A56 and the port B57, and the description is completed. In FIG. 17, the primary power input from the power unit meshes with a sprocket or gear 20, which in turn meshes with a sprocket or gear 42, which is mounted on shaft 62 and shaft 62 is a bearing. It is carried by 63 and then drives the curved axis variable discharge pump 69, which in turn supplies and receives hydraulic fluid via ports A60 and B61.

The lens 64 of the bend axis variable discharge pump is controlled from zero to the maximum flow rate by the rotation of the control screw 65, and then is continuously driven by a motor 66 driven via commands from the microprocessing control unit shown in FIG. Is controlled. Ports 60 and 61 of the variable-shaft variable discharge axial piston pump are substantially interconnected with ports 67 and 68 of the unitized differential reaction hydraulic / mechanical power transmission 58, and a piston rotating group 53.
The control of the reaction force and the speed can be performed by the action as described above.

As described with reference to FIG. 17, the present invention is fixed to the commutator plate 51,
The piston group 53 and the shaft A-
It can also be used with a variable axis pump pump unit 69 operatively connected to a similar arrangement of cam tracks 40 carried for rotation about the shaft of A, the supporting members of which are also described above. Axis A-A with unit 58 connected to bent axis pump assembly 69 via offset port
, But both units are operably connected to the power transmission unit 10 in the same manner as described in FIG. 13 as being independent of the power transmission unit 10. In a further embodiment of the present invention, when using the above method to provide a rotating piston group, it acts against the cam plate in the form of an axial or radial piston to achieve the same effect as described. With flat-sided rollers, with axle-curved sided rollers, with axles or balls or rollers, or without an axle (shown later in FIGS. 23 and 24 for one of the embodiments) or Acts against any of the matching closed cam tracks at any of the radial positions.

The invention is suitable for various embodiments for all types of vehicles, trucks and tractors. In the case of tractors and trucks, in order to understand the requirements and the application of the present invention, there are exceptional torque requirements for both trucks and tractors and the extremes of external and highly variable energy demands. Exploring the currently available and most advanced technologies and methods to meet high load and extremely fluctuating torque requirements will extend the features and applications of the present invention to tractors, trucks and other heavy load transport and various machine modes. It is useful for better understanding. FIG. 18 shows the case where the first gear at the start of the 24-gear transmission is 6
As shown in FIG. 9, a forward speed of 3.4 km / h is obtained, and the engine repeats the cycle of each gear in each gear as the gear progresses, and each cycle starts at a low rotational speed of the engine. High torque load demands are while the engine is at low speed. This also results in high fuel consumption in terms of producing low power, and the conversion of the fuel-air mixture to power is less efficient. As the engine speed increases, the power output increases until the engine speed and torque balance. This point is usually located at the center (mid range) of each gear, and is the point at which fuel consumption is most efficient. As the engine continues to increase in speed, fuel consumption increases and fuel conversion efficiency to power decreases in special gears. First
In FIG. 8, this cycle is repeated as represented graphically through the 24th gear, and the engine cycles as described above. Curve 70 represents the desired average gear ratio, and as a variable, the engine operates in a constant balance between engine speed and torque load, at which point fuel consumption is lowest and engine efficiency converts fuel to power. It is the biggest to do. It is an object of the present invention to always operate the engine in this balanced state for maximum efficiency.

FIG. 19 is a diagram of a composite power train required to achieve the 24-speed transmission shown in FIG.

The power train diagram of FIG. 19 begins with the engine 71. Engine torque and horsepower (power) are transmitted through torque converter 72 to smooth the impact loads that occur when gear changes and clutches are switched either manually or automatically, and torque converters are particularly low speed, high load. Used at startup. An efficient aspect of the energy transfer of a torque converter is that it also causes power loss to be dissipated as heat due to energy slip. As shown in FIG. 19, the tractor requires a power shift reverse set 73, a driving clutch 74, a six-speed traveling gear 75, and a crawler gear / working gear 76 in addition to the 24-speed forward gear.

FIG. 20 is a schematic cut-away view of a 24-gear transmission showing a complex and expensive transmission achieved with a 24-gear or gear ratio that smoothes the power performance curve to an acceptable level. . Stepless gear ratio with minimal moving parts, low power while increasing powertrain efficiency, reducing shock loads, eliminating engine gasping or overspeed, while improving fuel economy and reducing pollution It is an object of the present invention to achieve the transformation.

FIG. 20 shows a turbo clutch / torque converter 76, a power shift reverse set 77, a drive clutch 78, a six-speed running gear 79, a crawler gear / working gear 80, and a four-wheel drive output 81 to a front axle.

FIG. 21 is an illustrative cutaway view of the turbo clutch / torque converter 76.

FIG. 22 is a graph of the effects of engine speed and torque balanced from the point of view of fuel efficiency. Baseline 82 indicates the actual speed range of the diesel engine as used in tractor applications. The torque level is shown in units of Newton meter (Nm) 83, and the torque graph is shown by 84. The power is shown in units of kilowatts (KW) 85 and the power curve is shown at 86. The effect of the engine speed and torque on fuel consumption can be seen as a fuel consumption curve 87 on the graph. 640N when the engine is at low rpm 88 at 1,000 rpm
It will be noted that at a high torque of m, the fuel consumption is 215 g / kwh. When the engine speed is 89 between 1,400 rpm and 1,600 rpm, the torque is 650 Nm, the fuel consumption is 92 and the lowest is 195 g / kwh, and when the engine speed is 90 and increases to 2,200 rpm, Torque is 85Nm at 91
To decrease. Fuel consumption increases to 225 g / kwh at 92.

It is an object of the present invention to operate the engine in the maximum efficiency range despite constantly fluctuating load and power requirements, which is shown at 1,450 rpm
It can be seen that this is the point at which the rotational speed and the torque are balanced for the specific engine in which the torque is 640 Nm and the fuel consumption is 195 g / kwh.

The graph of FIG. 22 is optimally selected in the highly fluctuating rotation speed and torque range provided by the present invention, and the input is continuously and reactively changed. As shown in FIG. It is only used to illustrate the control of the output from the microprocessing control unit.

FIG. 23 is an embodiment of the invention suitable for high torque, high load vehicles, especially heavy traction by trucks and high torque workloads in tractor forward and reverse travel.

The main transmission 11 is sized according to the high load and torque requirements, but embodies the same principle as described in FIG. 9, with the power transmission unit 10 having two inputs, the main transmission 11 and the differential gear assembly 9. Are combined to form The speed fluctuation required between the bevel gear 5 is mechanically supplied with power from the primary power input drive line and the annular pinion carrier 13, and the same reaction force is generated as described in FIGS. 4, 5, 6 and 7. And control.

FIG. 23 shows a compact integrated combination of a hydraulic / mechanical differential power control unit as part of the present invention, integrating the same power transmission 10 as described in FIG. To control the speed fluctuation and reaction force.

FIG. 23 particularly illustrates the compact features and high torque characteristics of this embodiment of the present invention. The primary power input drive line 42 is mounted on an annular support member 92 which in turn supports and fixes an annular internal convolut cam 93. A partially cut end view of the cam track and axial piston assembly is shown in FIG.
It is shown in FIG. In FIG. 23, the cam 93 is fixed to the ported valve body 94 and rotates together with the cam 93, so that the hydraulic oil galleries 95 and 96 are correctly aligned regardless of the rotational position to supply high pressure. The oil flow is returned and properly positioned with respect to the cam hole 97 shown in FIG.

In FIG. 23, the stationary housing 98 is mounted on the non-rotating part of the valve body 99 and interconnected by parts 100 and 101 to the corresponding separate internal gallery and communicates with the rotating piston group 102, This causes high pressure and return oil to be pumped into the cylinder chambers 103 (a), 103 (b) at the correct point in time, extending the pistons 104 (a), 104 (b), and the rollers 105 (a), 105 (b) To engage the abutment at the appropriate position of the truck, to connect the linear thrust of the pistons 104 (a), 104 (b) to a rotational movement and / or to control the rotational speed of the annular pinion carrier 13 relative to the bevel gear 5 As a restraining unit. This includes a linear gallery communicating with the corresponding cylinder chambers 103 (a), 103 (b), which are linearly and radially separated, and are achieved by the galleries 95, 96 while the other cylinders are radially separated. It is arranged and evenly arranged in the rotating piston group, and a sectional view of this part is shown in FIG.

By the action of a rotating piston group 102 mounted directly on an annular pinion carrier 13 rotatably carrying a planetary bevel pinion gear 7 arranged therein, a change in output speed and a torque force are caused, It is restrained by the rotating piston group 102 and acts against the cam track 93 in a hydraulic differential manner as described above.

In FIG. 23, the hydraulic differential speed control is performed on the axial piston roller 105 (a
), 105 (b) interact with the cam track 93 on the outer periphery of the power transmission powertrain assembly 106 for maximum efficiency and maximum torque application.

The dynamic hydraulic differential power control is located on one side of the annular pinion carrier 13 and is compactly integrated with the power transmission unit 10 to form an integrated assembly.

FIG. 23 shows a compact, high-torque, efficient configuration suitable for directly retrofitting differentials to vehicles, trucks and tractors for all applications described, from reverse to overdrive. The present invention is shown to function both as a continuously variable transmission that can be varied to infinity and as one or both to provide differential energy to both axles 18,19. In one embodiment (not shown), differential 9 may incorporate a limited slip or controllable differential lock assembly. The present invention, as explained in FIG.
The universal and propeller shafts of the wheel drive vehicle or tractor are connected to the front and rear differentials and axle assembly to provide half axles 18, 1
It can also be advantageously located at the midpoint of a four-wheel drive tractor with nines.

FIG. 24 shows the transmission to the bevel gear 5 through the pinion 107 and the bevel gear 108 and to the gear or sprocket 20 via the gear and sprocket 42, which is then mounted on the bend axis variable discharge pump 69, This in turn supplies and receives hydraulic fluid via ports 60 and 61, and port 100, within integrated differential reaction hydraulic / mechanical power transmission 106, as described in FIG.
FIG. 23 is an embodiment of the invention described in FIG. 23 with a primary mechanical drive line properly connected to ports 60 and 61 properly connected to 101. As described in FIG. 24, the present invention is particularly suited for retrofitting transmissions and differentials, and is located in a standard position for vehicles and especially truck and tractor differentials, and has a built-in A Only the high-speed tail shaft input at the point is required, the point A is connected to the pinion 107, and the other external inputs are the micro-energized lead screw 66 described in FIG. 14 on the motor energized lead screw 66 as described in FIG. This is a control device based on an input from the processing control unit 45.

All the differential hydraulic / mechanical power control units already mentioned will undergo a reversal of the reaction force, as described in connection with the invention in all figures, and, where appropriate, To prevent cavitation, a check valve capable of drawing hydraulic fluid will be installed on the low pressure side of the hydraulic circuit. If an energy reversal occurs, the check valves on both sides of the circuit may be integrated with a pressure relief valve and / or a fill pump to supply low pressure hydraulic fluid to the alternately low pressure side of the circuit.

FIG. 25 functions as described in FIG. In FIG. 25, the fan-shaped portion of the cam track 93 has a rotating piston group 10 as viewed from the end face.
2. Cylinder chamber 103 (a), 103 (b), 103 (c), piston 1
04 (a), 104 (b), 104 (c), rollers 105 (a), 105 (b)
, 105 (c). The cam track 93 has oppositely shaped tracks, which are symmetrical and therefore balanced. Piston 10
4 (a), 104 (b), 104 (c) and other pistons are radially disposed and actuated to maintain rotational balance. The radial arrangement of the cylinders and pistons of the radial piston rotating group 102 is positioned to correspond to the corrugated shape of the cam track 93 while the piston 104 (c)
And the roller 105 (c) enter the position of energy transmission either in or out of the wave shape of the cam track 93, and the roller 105 (b) and the piston 104
(B) completes the cycle and cylinder chamber 103 (b) is at maximum discharge.
This action smoothly transitions from rotation to reciprocation of the oil rod, in a stepwise manner, over the complete circumference of the assembly in a stepwise manner over multiple pistons and corrugations, as shown in FIG. Bending axis variable discharge axial piston pump 69 shown
Is continuously transferred and controlled by

FIG. 26 shows a schematic cut-away view of a four-wheel drive tractor for one embodiment of the present invention 106 as described in FIG. 17, which preferably differs from the front and rear axles 108 and 109. Conveniently mounted to provide dynamic power.

FIG. 27 is a diagrammatic cutaway view of a tandem drive assembly 110 that replaces the main transmission / gearbox and primary drive differential for one embodiment of the present invention 111 as described in FIGS. 23 and 24. Is shown.

The mechanical program shaft from the in-line engine provides power and
, And the engine mechanical input power is in the direction of arrow 112. The differential differential housing mounted with the transmission assembly is indicated by arrow 113, the transmission differential full axle is connected at 114 and 115, and if in place both axles 116, 1
17 would have provided a power output. The integrated power unit 111 is the second
It will be appreciated that it can be used as a tandem and / or single axle machine as described in FIG.

FIG. 28 shows an embodiment of the present invention. As shown in FIG. 9, the outer main transmission 11 has an annular pinion carrier 13 disposed on the outside and rotatably supporting the planetary bevel pinion gear 7. It is also possible to have the differential gear assembly 9 mounted on one side of the outer main transmission 11 with By this means, the transmission unit 10 with very high torque power can be
The maximum diameter of the main transmission 11 including the first bevel gear 5 and the planetary assembly 12 arranged coaxially to rotate about the axis AA can be reduced to an acceptable outer diameter. The planetary gear assembly 12 is mounted internally and is carried on an internal annular pinion carrier 13. The spindle 4 projects radially and is fixed to a hollow annular pinion carrier 13. The spindle 4 carries a bevel pinion gear 7, which is freely rotated around the spindle 4 by a bearing 8. A gear or sprocket 43 is attached to the outer end of the spindle 4. The functions of the transmission unit 10 are described in FIGS. 8 and 9, comprising a primary drive line shown in FIG. 28 as A, a secondary variable drive line 43 shown as B, and a second bevel gear 6 shown as C. The power output is then fixed at D directly to the differential carrier 20.

As a result, the output drive of the main transmission 11 via the second bevel gear 6 is transmitted to one side of the main transmission 11 and transmitted to the large-diameter differential gear assembly 9 mounted on one side, and the bevel gear 5, The diameter can be equal to 6 and can accept very high torque power standard differential transmissions.

The standard differential function supplies power to the axle 19 via the side gear 17, the hollow bevel gear 6 via the side gear 16, and the carrier assembly 118.
Power is supplied by the axle 18 through the bevel gear 5.

The differential can be arranged on the left or on the right, and in principle the ninth, twelfth, thirteenth,
14, 15, 16, 17, 23, 24, 26 It will be appreciated that various embodiments may be incorporated as shown and described in FIG. The invention described in FIG. 28 is particularly suitable for the application of the embodiment described in FIGS. 24 and 27, in particular for the direct replacement of standard differentials, as shown in FIG. When used in extremely high torque applications, such as heavy duty vehicles, it thereby accepts a reduction in diameter for ground clearance within both narrower radial spaces with both gearbox and differential functions. It is only a matter of scale to use the same technology as described in FIG. 28, and the main provision of a continuously variable shift as well as a differential for all types of vehicles, trucks and tractors as an integrated embedded powertrain It will be appreciated that the described embodiments of the hydraulically controlled hydraulic / mechanical differential power control unit with power transmission unit and side mounted differential are combined.

Although the present invention has been described with particular reference to the preferred embodiments, it is to be understood that the present invention may be embodied in various alterations and modifications that differ particularly from those set forth in the preceding specification. is there. These changes and revisions are possible without departing from the scope of the present invention.

[Brief description of the drawings]

FIG. 1 is a graph of the efficiency of an internal combustion engine in relation to changes in torque and engine speed.

FIG. 2 is a shift table showing a power pattern of the automatic transmission for a vehicle, showing steps of power input and vehicle speed.

FIG. 3 is a graph showing a comparison between a conventional multi-geared mechanical transmission and a powertrain transmission of the present invention using a reaction force controlled by a confinement oil in the form of a reverse hydraulic transmission on a secondary drive line.

FIG. 4 is a diagram of the ratio and reaction force associated with the power output of the primary and secondary input power drive lines and the outer transmission.

FIG. 5 is a diagram of the ratio and reaction force associated with the power output of the primary and secondary input power drive lines and the outer transmission.

FIG. 6 is a diagram of the ratio and reaction force associated with the power output of the primary and secondary input power drive lines and the outer transmission.

7A and 7B are a diagram and a partial cross-sectional schematic diagram showing a linear shape converted to a rotary shape of an outer transmission.

FIG. 8 is a schematic partial cross-sectional view of an embodiment of the power transmission unit according to the first aspect of the present invention, including the relationship and effects of the outer and inner transmissions and the primary and secondary power input drive lines.

9 is a cross-sectional view of an actual embodiment of the power transmission unit shown in FIG. 8 showing a relationship between a primary power drive line from a single power unit and a secondary variable power drive line.

FIG. 10 is a schematic partial cross-sectional view of an embodiment of a differential displacement hydraulic / mechanical variable secondary power drive line reactively connected and controlled to an outer power transmission indicating a maximum discharge position.

FIG. 11 is a partial cross-sectional view of the secondary power drive line shown in FIG. 10 at a zero displacement position.

FIG. 12 shows the combined control of the ratio and reaction force and the combined output of the secondary power drive line shown in FIGS. 10 and 11 in which the full range of discharge and the primary and secondary power drive lines are directly connected to the outer power transmission. It is a partial sectional view.

FIG. 13 is a cut-away view of an actual embodiment of the power transmission, partially illustrated in FIG. 12, including front wheel drive including direct replacement of the differential and gearbox when located in a conventional position; , Rear-wheel drive and differential displacement dynamic hydraulic / mechanical variable secondary power drive line and primary power drive shown in FIGS. 10, 11, and 12 suitable for general placement and application of 4WD vehicles, tractors, and trucks One embodiment of the line interrelation is differential in front and rear axle driven vehicles, trucks and tractors.

FIG. 14 relates to a gasoline or diesel engine, a primary and secondary power line control shown in FIG. 12, and an engine optimally arranged for a horizontal engine mounted on a front wheel drive vehicle of the power transmission shown in FIG. FIG. 4 is a partial cross-sectional diagram of a gasoline and diesel engine embodiment of a powertrain according to a second aspect of the present invention showing the interrelationship in the preferred embodiment with power output.

FIG. 15 is a partial cross-sectional diagram of a simplified embodiment of the powertrain shown in FIG. 14 suitable for rear power output of a traditional linear engine arrangement in vehicles, trucks and tractors.

FIG. 16 shows a power transmission as shown in FIG. 13 which combines one embodiment of a dynamic hydraulic / mechanical variable secondary power drive line integrating the outer transmission and inner differential and a differential displacement of the primary drive line. FIG. 4 is a sectional view of an actual embodiment of the unit.

17 is a partial sectional view of the embodiment of the power transmission shown in FIG. 16 in which the primary and secondary drive lines are integrated.

FIG. 18 is a graph of selected gear versus forward speed for a prior art 24-speed gear transmission.

FIG. 19 is a diagram of the prior art powertrain required to achieve the 24-gear ratio shown in FIG. 18;

FIG. 20 is a cutaway view of a 24 stage gear transmission of the type shown in FIG. 19;

21 is a partial cutaway view of a type of torque converter used in the transmission of FIG. 20.

FIG. 22 is a graph showing changes in fuel consumption, torque and power with respect to engine speed for a typical diesel engine.

FIG. 23 illustrates an embodiment of the powertrain of the present invention suitable for high torque loads.

FIG. 24 shows an application of the embodiment of the invention shown in FIG.

FIG. 25 shows the working part of the arrangement shown in FIGS. 23 and 24.

FIG. 26 shows a partially cutaway view of a four-wheel drive tractor incorporating a powertrain according to an embodiment of the present invention.

FIG. 27 shows a partially cutaway view of a tandem drive assembly incorporating a powertrain according to the present invention.

FIG. 28 shows a further embodiment of the powertrain according to the invention.

──────────────────────────────────────────────────続 き Continuation of front page (81) Designated country EP (AT, BE, CH, CY, DE, DK, ES, FI, FR, GB, GR, IE, IT, LU, MC, NL, PT, SE ), OA (BF, BJ, CF, CG, CI, CM, GA, GN, GW, ML, MR, NE, SN, TD, TG), AP (GH, GM, KE, LS, MW, SD, SL, SZ, UG, ZW), EA (AM, AZ, BY, KG, KZ, MD, RU, TJ, TM), AE, AL, AM, AT, AU, AZ, BA, BB, BG, BR , BY, CA, CH, CN, CU, CZ, DE, DK, EE, ES, FI, GB, GD, GE, GH, GM, HR, HU, ID, IL, IN, IS , JP, KE, KG, KP, KR, KZ, LC, LK, LR, LS, LT, LU, LV, MD, MG, MK, MN, MW, MX, NO, NZ, PL, PT, RO, RU, SD, SE, SG, SI, SK, SL, TJ, TM, TR, TT, UA, UG, US, UZ, VN, YU, ZA, ZW The rotation speed of the second rotating element (28) is determined according to the rotation of the first rotating element (22).

Claims (17)

[Claims] 1. A power split device having two rotational outputs having a variable relative rotational speed from a single rotational input, wherein the device is rotationally driven about a rotational axis and has a first rotational output.
A first rotating element having a rotational output, a second rotating element rotatable about the axis of rotation and having a second rotational output, a first fluid chamber associated with the first rotating element, and the first rotating element First regulating means for changing the volume of the first chamber in response to rotation of the second rotating element, and a second regulating means associated with the second rotating element.
Fluid chamber; second regulating means for changing the volume of the second chamber in response to rotation of the second rotating element; and between the first and second chambers during rotation of the first and second rotating elements. Commutator means for at least regularly establishing a closed fluid flow transmission, wherein the relative timing of changes in the volumes of the first and second chambers is responsive to rotation of the first rotating element. A power split device that determines the rotational speed of a vehicle. 2. The apparatus according to claim 1, wherein one of said first or second regulating means is selectively adjustable for adjusting a rotational speed of said second rotating element. 3. The first and second regulating means include corresponding first and second pistons movable within the first and second fluid chambers, respectively, to change corresponding volumes of the chamber. 3. The device according to 1 or 2. A third fluid chamber associated with the first rotating element; third regulating means for changing a volume of the third chamber in accordance with rotation of the first rotating element; An associated fourth fluid chamber and the second fluid chamber;
Fourth regulating means for changing the volume of the fourth chamber in response to rotation of a rotating element, and wherein a closed fluid flow between the first and third chamber sets and the second and fourth chambers And the commutator means for selectively establishing the communication of the device. 5. The first and third chambers are substantially radially opposed across the axis of rotation, and the second and fourth chambers are substantially radially across the axis of rotation. 5. The device according to claim 4, wherein the device faces. 6. The piston is movable in a direction generally parallel to the rotation axis,
4. The system according to claim 3, further comprising first and second swash plates for operating said piston.
An apparatus according to claim 1. 7. The method according to claim 7, wherein at least one of said swashplates comprises said first or second swashplate.
7. The apparatus of claim 6, wherein the apparatus is selectively adjustable to control a change in a corresponding one of the chambers. 8. The apparatus according to claim 1, wherein said commutator means includes one or more openings substantially circumferentially surrounding said axis of rotation. 9. The apparatus according to claim 8, wherein said openings are kidney bean shaped. 10. The apparatus according to claim 1, wherein said first rotating element comprises a hollow cylindrical housing at least partially surrounding said second rotating element. 11. The apparatus of claim 10, wherein said first fluid chamber, said commutator means and said second fluid chamber are contained within said housing. 12. The apparatus of claim 11, wherein said first fluid chamber is fixed against rotation, said commutator means rotates with said housing, and said second fluid chamber rotates with said second rotating element. . 13. The timing of a change in the volume of the plurality of chambers is the first timing.
And wherein the second rotating element is adjustable for rotation in conjunction with one another.
An apparatus according to any one of claims 1 to 12. 14. The timing of the change in the volume of the plurality of chambers is the first timing.
14. Apparatus according to any of the preceding claims, wherein the second rotating element can be adjusted to rotate at half the rotational speed of the rotating element. 15. The fluid chamber according to claim 1, wherein said fluid chamber is filled with oil.
An apparatus according to any one of the preceding claims. 16. A power split device, comprising an outer main transmission and an inner differential gear assembly, wherein the power split device has two rotational outputs having variable relative rotational speeds, the device being driven around a rotating shaft by a power unit. A first rotating element having a first rotating output, a second rotating element rotatable about the rotation axis and having a second rotating output, a first fluid chamber associated with the first rotating element, First regulating means for changing the volume of the first chamber in response to rotation of the one rotating element, a second fluid chamber associated with the second rotating element, and the second fluid element in response to rotation of the second rotating element. Second regulating means for changing the volume of the second chamber; and first and second rotating means for rotating the first and second rotating elements. Commutator means for at least regularly establishing a closed fluid flow transfer between the first and second chambers, wherein the relative timing of the change in volume of the first and second chambers is dependent upon the rotation of the first rotating element. A power transmission unit for determining a rotation speed of the second rotary element in accordance with the power transmission unit, wherein the main transmission has two rotation input means driven by a first rotation output and a second rotation output of the power split device, respectively. The two input means are rotatably connected to the rotary output means, whereby the rotational speed of the output means changes in proportion to the algebraic means of the rotational speed of the two input means, and the differential gear assembly is located inside the main transmission. Having rotational input means operably connected to the two differential rotational output means, Input means of the output means and the differential gear assembly of the main transmission is operably connected, the power transmission unit. 17. A vehicle power train that can be continuously controlled over a set operating range, including a single power unit, a power split device, an outer main transmission, and a power transmission unit including an inner differential gear assembly. The power split device provides two rotational outputs having a variable relative rotational speed, the power split device comprising a first rotary element having a first rotational output driven about a rotation axis by the power unit; A second rotating element rotatable about an axis and having a second rotating output, a first fluid chamber associated with the first rotating element, and a volume of the first chamber responsive to rotation of the first rotating element. A first regulating means for altering and a second flow associated with the second rotating element. A second regulating means for changing the volume of the second chamber in response to the rotation of the second rotating element; and a second regulating means for rotating the first and second rotating elements during rotation of the first and second rotating elements. Commutator means for at least regularly establishing a closed fluid flow transmission, the relative timing of the change in the volume of the first and second chambers being dependent on the rotation of the first rotary element. The main transmission has two rotation input means which are respectively driven by a first rotation output and a second rotation output of the power split device, and the two input means serve as rotation output means. Operably connected so that the rotational speed of the output means varies in proportion to the algebraic means of the rotational speeds of the two input means, The gear assembly has rotational input means disposed within the main transmission and operatively connected to two different rotational output means, wherein the output means of the main transmission and the input of the differential gear assembly. A vehicle powertrain, wherein the means is operably connected.